DE3880864T2 - Active suspension system with anti-roll control device. - Google Patents

Description

The present invention relates to an anti-roll suspension control system for motor vehicles, comprising a suspension system between a vehicle body and a suspension member rotatably supporting a vehicle wheel, the suspension system comprising a variable pressure pressure chamber filled with a working fluid of controlled pressure, a pressure adjusting device in communication with the variable pressure pressure chamber for adjusting the pressure of the working fluid in the pressure chamber, a sensor for monitoring the roll conditions of the vehicle and for generating a signal representative of the roll intensity, a control unit which receives the signal representative of the roll intensity, generates an anti-roll suspension control signal based on the signal representative of the roll intensity and controls the pressure adjusting device for adjusting the suspension characteristics for suppressing roll of the vehicle, and a vehicle speed sensor which detects a signal representative of the vehicle speed. representative signal is generated.

Such an anti-roll suspension control system is known from EP-A-0 114 757. In this document a vehicle suspension system is disclosed comprising suspension units and a control unit which receives inputs from transducers connected to these suspension units, inputs from accelerometers in response to longitudinal and lateral accelerations, an input from a speedometer in response to a forward speed and an input from a transducer in response to a hydraulic supply pressure. An analog processor is connected to this control unit for generating output signals which ultimately cause the adjustment of an actuator integrated in this suspension system to achieve a desired vehicle attitude.

U.S. Patent 4,702,490, issued October 27, 1987 to Hirotsugu Yamaguchi et al. and assigned to the assignee of the present invention, discloses an actively or positively controlled automotive suspension system including hydraulic cylinders for generating a variable damping force against relative movements of a vehicle suspension member and a vehicle body in the compression and rebound directions for adjusting the suspension characteristics and for adjusting the vehicle body attitude.

A similar actively controlled suspension system is disclosed in copending U.S. Patent Application 060,911, filed June 12, 1987, and assigned to the assignee of the present invention; and in EP-A-0 249 246 and EP-A-0 249 209, which correspond respectively to copending U.S. Patent Applications 061,368, filed June 15, 1987, and 059,888, filed June 9, 1987, both of which are assigned to the assignee of the present invention. The latter three inventions include anti-roll suspension control features for adjusting damping characteristics at the respective vehicle wheels to suppress vehicle roll.

DE-A-3 408 292 discloses a vehicle suspension system provided, inter alia, with an anti-roll control system. The document refers to the earlier DE-A-3 212 433, which is said to disclose an active vehicle suspension system controlling the relative positions of the vehicle structure and the wheels. An active suspension system according to DE-A-2 441 172 senses the positional changes of the wheels or suspension members with respect to the vehicle structure as well as vertical accelerations to control the damping force of the system. It is stated that similar prior art is disclosed by DE-A-3 101 194.

The inventors of DE-A-3 408 292 intend to create an active suspension control in which both the vertical movement of the vehicle body and its rotational movement with respect to the longitudinal and transverse axes can be controlled independently.

In such an active control suspension system, an anti-roll suspension control is generally implemented using the applied lateral acceleration. To satisfactorily suppress vehicle roll, it is necessary to provide sufficient gain for a control loop dependent on the lateral acceleration. On the other hand, however, a larger gain in the control loop tends to increase the magnitude of self-induced vibrations in the roll direction. Therefore, to minimize or maintain the magnitude of the self-induced vibrations at an acceptable level, the gain of the control loop must be limited. The limited loop gain in the control loop tends to make anti-roll characteristics of the suspension system unsatisfactory.

On the other hand, the inventor of the present invention has noticed that uncomfortable self-induced vibrations tend to occur in a range of fairly low vehicle speeds. It has also been found that at higher vehicle speeds, the magnitude of self-induced vibrations becomes smaller or self-induced vibrations do not impair ride comfort.

It is therefore an object of the invention to provide an actively controlled suspension system whose control characteristics are variable according to the vehicle speed.

Another object of the invention is to provide an anti-roll control system which is variable according to the vehicle speed.

The above-mentioned and other objects are achieved by an active control suspension system according to the invention, which is characterized in that the control unit (100) receives the signal representative of the vehicle speed and adjusts the response characteristics of the anti-roll suspension control in accordance with the signal representative of the roll intensity in dependence on the signal value representative of the vehicle speed, so that the response characteristic is adjusted to be lower than a predetermined first value according to a predetermined response characteristic curve which is substantially proportional to the signal representative of the vehicle speed, so that self-induced vibrations are suppressed which are caused by the anti-roll suspension control in a roll direction occur while the signal value representative of the vehicle speed is less than a predetermined vehicle speed, and the response characteristic is maintained at or near the first value so that the anti-roll effect is enhanced while the signal value representative of the vehicle speed is greater than or equal to the predetermined vehicle speed.

In practical control, the gain in a region of vehicle speed where the larger loop gain of the control loop causes significant and unacceptable levels of self-induced lateral vibrations is changed to be smaller than that in another vehicle speed range. Namely, since the larger gain in a region of fairly low vehicle speeds may cause self-induced lateral vibrations, the gain may be set to a smaller value as long as the vehicle speed is lower than a predetermined value.

According to one embodiment of the invention, an anti-roll suspension control system for a motor vehicle comprises a suspension system between a vehicle body and a suspension member rotatably supporting a vehicle wheel, the suspension system comprising a variable pressure pressure chamber filled with a working fluid at a controlled pressure, and pressure adjusting means in communication with the variable pressure pressure chamber for adjusting the pressure of the working fluid in the pressure chamber, a first sensor for monitoring the vehicle speed which generates a vehicle speed signal representative of the vehicle speed; a second sensor for monitoring the vehicle roll condition which generates a signal representative of the roll amount, and a control unit which receives the vehicle speed signal and the signal representative of the roll amount and, on the basis of the signal representative of the roll amount, generates an anti-roll suspension control signal for controlling the pressure adjusting device for adjusting the suspension characteristics for suppressing the vehicle roll, which control unit adjusts the response characteristic of the anti-roll suspension control in response to the signal representative of the roll amount which depends on the vehicle speed signal.

The control unit determines that the value of the vehicle speed signal is less than a predetermined value and then lowers the response characteristics. The control unit maintains the response characteristics at a predetermined constant first value as long as the value of the vehicle speed signal is greater than or equal to the predetermined value. The control unit changes the response characteristics between the first value and the value 0 according to the signal value representative of the vehicle speed while the signal value representative of the vehicle speed is lower than the given value. The control unit sets the given value to a value representative of a fairly low speed.

The pressure adjusting device comprises a source of pressurized fluid connected to the variable pressure pressure chamber and a pressure control valve including an electrically operable actuator, the actuator connected to the control unit and receiving the anti-roll suspension control signal and operating the pressure control valve to adjust the amount of working fluid introduced and discharged with respect to the variable pressure pressure chamber and thereby adjust the damping characteristics of the suspension system. The control unit comprises a gain-controlled amplifier which determines the response characteristics and adjusts the gain of the gain-controlled amplifier in accordance with the signal value representative of the vehicle speed. The control unit compares the signal value representative of the vehicle speed with a predetermined low vehicle speed criterion to maintain the gain of the gain-controlled amplifier at a constant first value when the signal value representative of the vehicle speed is greater than or equal to the low vehicle speed criterion and decreases the gain according to the signal value representative of the vehicle speed being less than the low vehicle speed criterion.

The control unit changes the gain of the gain-controlled amplifier in a range between the first value and the value 0 depending on the signal value representative of the vehicle speed. The control unit changes the gain of the amplifier whose gain is controlled linearly according to the signal value representative of the vehicle speed.

Short description of the drawings

In the drawings

Fig. 1 is a schematic representation of the general structure of an actively controlled suspension system;

Fig. 2 is a sectional view of a pressure control valve unit for use in the preferred embodiment of the active control suspension system according to the invention;

Figs. 3(a) and 3(b) are sectional views similar to Fig. 2, but showing the state in which a piston in the pressure control valve unit is pushed downward to reduce the fluid pressure in a fluid chamber of a hydraulic cylinder in the suspension system and upward to increase the fluid pressure in the fluid chamber;

Fig. 4 is a graph showing the change of fluid pressure in a hydraulic cylinder controlled by the pressure control valve unit of Fig. 2;

Fig. 5 is a schematic and diagrammatic representation of a hydraulic circuit for use in the preferred embodiment of the line pressure control according to the invention;

Fig. 6 is a schematic block diagram of the preferred embodiment of the suspension control system of Fig. 1;

Fig. 7 is a diagram showing the relationship between a gain control signal R and a gain degree Ky used in the suspension control system of Fig. 6 in a gain controlled amplifier which sets roll suppressing suspension control characteristics;

Fig. 8 is a diagram showing the relationship between the gain control signal R and a vehicle speed;

Fig. 9 is a graph showing the relationship between the vehicle speed and the gain rate Ky;

Fig. 10 is a diagram showing the relationship of gain control signals Ef and Er and gain levels Kf and Kr used in the suspension control system of Fig. 6 in gain controlled amplifiers that cooperate with each other to set pitch suppressing suspension control characteristics;

Figs. 11(a) and 11(b) are diagrams showing the relationship between a longitudinal acceleration Gx and gains Ef and Er;

Fig. 12 is a diagram showing another example of a relationship between the longitudinal acceleration Gx and gains Kf and Kr;

Fig. 13 is a diagram showing a model for use in the discussion of the anti-roll suspension control logic;

Fig. 14 is a diagram showing stable and unstable regions of the roll control gain Ky;

Fig. 15 is a diagram showing change characteristics of a frequency transmission coefficient in one cycle of the anti-roll control;

Fig. 16 is a flow chart showing a routine for setting the roll suppressing suspension characteristics control gain; and

Fig. 17 is a flowchart showing a program for setting the pitch suppressing suspension characteristics control gain.

Description of the preferred embodiment

Referring to the drawings, particularly to Fig. 1, a vehicle has four suspension systems 11FL, 11FR, 11RL and 11RR for suspending the vehicle body 12 from front left, front right, rear left and rear right road wheels 14FL, 14FR, 14RL, 14RR, respectively. Each of the front left, front right, rear left and rear right suspension systems 11FL, 11FR, 11RL and 11RR includes a suspension member 13 such as a wishbone, a suspension link, etc., and a suspension unit 15 disposed between the vehicle body 12 and the suspension member 13. The suspension unit 15 has a hydraulic cylinder 15A, which serves as an actuator for generating a damping force against compression and rebound movements between the vehicle body and the suspension member, and a coil spring 16.

It should be noted that in the embodiment shown, the clock spring 16 is not necessary for damping the dynamic kinetic energy and is only intended to resiliently support static loads exerted between the vehicle body and the suspension member. It should of course still be possible to use the clock spring, which can be strong enough to dampen part of the dynamic kinetic energy during the relative compression and rebound movement of the vehicle body and the suspension member.

The hydraulic cylinder 15A has a hollow cylinder housing 15a filled with a viscous working fluid and a piston 15c which is tightly and slidably arranged in the interior of the cylinder housing and divides the cylinder space into upper and lower fluid chambers 15d and 15e. A piston rod 15b extends through one end of the cylinder housing 15a. The other end of the cylinder housing 15a is connected to either the vehicle body 12 or the suspension member 13. On the other hand, the end of the piston rod 15b is conversely connected to either the suspension member 13 or the vehicle body 12.

The hydraulic cylinder 15A of the suspension unit 15 is connected to a hydraulic pressure source unit 20 via a hydraulic circuit including a pressure control valve 18. The pressure control valve 18 is electrically operable and connected to a microprocessor-based control unit 100. The hydraulic circuit 19 includes a feed line 19s and a drain line 19d. The pressure source unit generally includes a pressure unit 20 and a reservoir tank 21. The pressure unit 20 is connected to the reservoir tank 21 for sucking the viscous working fluid in the reservoir tank 21 and supplying it to the pressure control valve 18 via the feed line 19s. On the other hand, the drain line 19d for returning the working fluid to the reservoir 21 is connected thereto.

As can be seen from Fig. 1, a pressure accumulator 22P is connected to the upper fluid chamber 15d of the hydraulic cylinder 15A via a pressure line 22B. A throttle valve 22V is inserted between the upper fluid chamber 15d and the pressure accumulator 22P. In the embodiment shown, the throttle valve 22V has a fixed throttle rate.

Fig. 2 shows the detailed structure of the hydraulic cylinder 15A and the pressure control valve 18. As can be seen from Fig. 2, the hollow cylinder housing 15a is provided with a connection port 15f which connects the upper fluid chamber 15d with an outlet port 18d of the pressure control valve 18 via a connection line 27. Although not clearly shown in Fig. 2, the lower fluid chamber 15e is formed as a closed space and filled with the viscous working fluid. The pressure of the working fluid in the lower fluid chamber 15e at an initial position of the piston 15c serves as a reference pressure and per se serves as a resistance for the downward movement of the piston.

The pressure control valve 18 has a valve housing 18A having the above-mentioned outlet port 18d, an inlet port 18b and a drain port 18c. The inlet port 18b, the drain port 18c and the outlet port 18d are each connected to a valve bore 18a formed inside the valve housing 18A. A spool 19 is slidably arranged in the valve bore 18a. The spool 19 has first, second and third collars 19a, 19b and 19c. As can be seen from Fig. 2, the third collar 19c has a smaller diameter than the first and second collars 19a and 19b. The third collar 19c defines a fifth pressure control chamber 18h, which is connected to the discharge opening 18c via a discharge channel 18f. An actuating piston 22c is also arranged in the valve bore 18a. The actuating piston 22c is spaced from the second collar 19b and thus defines a second pressure control chamber 18i, which is connected to the discharge opening 18c via a discharge channel 18e. An annular pressure chamber 18j is formed between the first and second collars 19a and 19b. The pressure chamber 18j is constantly connected to the outlet opening 18d and thereby also to the upper fluid chamber 15d. On the other hand, the pressure chamber 18j shifts according to the shift of the control slide 19 for selective communication with the inlet opening 18b and with the discharge opening 18c. On the other hand, a pressure control chamber 18k is formed between the first and third collars 19a and 19c. The pressure control chamber 18k is connected to the outlet opening 18d via a pilot channel 18g. A preload spring 22d is inserted between the actuating piston 22c and the control slide 19. The actuating piston 22c contacts an actuating rod 22a of an electrically actuable actuator 22, which comprises an electromagnetic cylinder coil. The electromagnet 22 forms a metering electromagnet.

To increase the feed pressure of the working fluid, the spool valve 19 is shifted to the position shown in Fig. 3(A) so that the pipe area at a throttle formed at the inner end of the inlet port 18b is increased by means of the collar 19a of the spool valve 19. On the other hand, to reduce the feed pressure of the working fluid, the spool valve is shifted to the position shown in Fig. 3(B) so that the pipe area at the throttle of the inner end of the inlet port 18b is reduced and opens the drain port 18 which is normally blocked by the collar 19b of the spool valve.

The structure of the pressure control valves should not be limited to the structure shown in Fig. 2, 3(A) and 3(B) but may be replaced by any suitable structure. For example, the pressure control valve structures shown in the above-mentioned EP-A-0 193 124 may also be used. The contents of the above-mentioned EP-A-0 193 124 are referred to as supplementing the present description.

As shown in Fig. 2, the metering solenoid 22 includes the operating rod 22a and a solenoid 22b. The solenoid 22b is energized by a suspension control signal V3 from the controller 100. In the illustrated embodiment of the pressure control valve, the working fluid pressure P at the outlet port 18d is variable according to the characteristic shown in Fig. 4. Namely, when the control value V3 represented by the suspension control signal is zero, the pressure P at the outlet port 18 becomes a predetermined offset pressure P0. When the suspension control signal value increases to positive values, the fluid pressure P at the outlet port 18d increases with a predetermined metering gain K1. By increasing the suspension control value V3, the pressure P at the outlet port 18d increases with a predetermined metering gain K1. Namely, the operating rod 22a is driven downward in Fig. 2 by an amount in the direction of the position of Fig. 3(A) so that an increase in the fluid pressure with the predetermined metering gain K₁ is achieved. The fluid pressure P at the outlet port 18d saturates at the output pressure P₂ of the pressure unit 20. On the other hand, when the suspension control signal value V₃ decreases, the pressure P decreases to zero by displacing the operating rod 22a in the direction of Fig. 3(B).

The actuating rod 22a of the metering solenoid 22 cooperates with the actuating piston 22c. Contact between the actuating rod 22a and the actuating piston 22c can be maintained by the spring force of the biasing spring 22d, which normally biases the actuating piston toward the actuating rod. On the other hand, the spring force of the biasing spring 22 is also exerted on the control slide 19, so that it is constantly biased downward in Fig. 2. The control slide 19 also receives an upward hydraulic force from the pressure control chamber 18k. Therefore, the control slide 19 assumes the position in the valve bore in which the downward biasing force of the biasing spring 22d coincides with the upward hydraulic force of the pressure control chamber 18k.

When compression occurs on the suspension member, the piston 15c of the hydraulic cylinder 15A moves upward and causes the rise of the fluid pressure in the upper chamber 15d. This causes the rise of the fluid pressure at the outlet port 18d of the pressure control valve 18. As a result, the fluid pressure in the pressure control valve 18k rises by the pressure introduced through the pilot passage 18g, so that the balance between the downward biasing force of the biasing spring 22d and the upward hydraulic force of the pressure control chamber 18k is destroyed. This causes the spool 19 to move upward against the spring force of the biasing spring 22d, as shown in Fig. 3(B). As a result, the passage area of the drain port 18c increases and the inlet port 18b is blocked. Therefore, the fluid pressure in the fluid chamber 15d is discharged through the drain port. Therefore, the increased fluid pressure in the fluid chamber 15d of the hydraulic cylinder 15A can be successfully absorbed so that the compression energy input from the suspension member is not transmitted to the vehicle body.

When a rebound movement occurs on the suspension member, the piston 15c of the hydraulic cylinder 15A moves downward and causes a decrease in the fluid pressure in the upper chamber 15d. This causes a decrease in the fluid pressure at the outlet port 18d of the pressure control valve 18. As a result, the fluid pressure in the pressure control chamber 18k decreases by the pressure introduced through the pilot passage 18g, so that the balance between the downward biasing force of the biasing spring 22d and the upward hydraulic force of the pressure control chamber 18k is destroyed. This causes a downward movement of the spool 19 against the spring force of the biasing spring 22d, as shown in Fig. 3(A). As a result, the passage area of the inlet port 18b decreases and the drain port 18c is blocked. Therefore, the fluid pressure in the fluid chamber 15d is increased by the pressure introduced through the inlet port. Therefore, the reduced fluid pressure in the fluid chamber 15d of the hydraulic cylinder 15A can be successfully absorbed, so that the rebound energy input from the suspension member is not transmitted to the vehicle body.

Since no flow resistance element, such as a throttle bore, a throttle valve, is placed between the fluid container 21 and the drain opening 18c, here, no damping force can be generated against the upward movement of the piston 15c in the hydraulic cylinder 15A in response to the compression movement of the suspension member. Since the damping force exerted on the piston 15c can serve to allow a part of compression energy to be transmitted to the vehicle body, thereby causing a rough ride, the embodiment of the suspension system shown can provide a satisfactorily high degree of ride comfort by completely absorbing the compression and rebound energy in the manner described above.

Fig. 5 shows a detailed circuit structure of the preferred embodiment of the pressure source unit for use in the suspension control system according to the invention. As described above, the pressure source unit comprises the pressure unit 20 which has a fluid pump and is connected to the tank 21 via a suction line 201 and which is driven by means of an automotive engine 200. The outlet of the pressure unit 20 through which the pressurized working fluid is passed is connected to the inlet port 18b of the pressure control valve 18 via the feed line 19s. A pressure regulating orifice 202 is arranged in the feed line 19s for suppressing a pulsating flow of the working fluid and regulating the output pressure of the pressure unit 20 to be supplied to the pressure control valve 18. A feedback line 19f is connected at one end to the upstream side of the pressure regulating orifice 202. The other end of the feedback line 19f is connected to the upstream side of the inlet of the pressure unit 20. Thereby, excess fluid between the pressure unit 20 and the orifice 202 is returned to the inlet side of the pressure unit.

An accumulator 203 is also connected to the feed line 19s and receives therefrom the pressurized fluid for storing the pressure. A one-way check valve 204 is arranged in the feed line 19s upstream of the junction between the accumulator 203 and the feed line 19s.

A relief line 205 is also connected at one end to the feed line 19s at a point between the pressure regulating throttle opening 202 and the one-way check valve 204. The other end of the relief line 205 is connected to the discharge line 19d. A pressure relief valve 206 is arranged in the relief line 205. The pressure relief valve 206 is responsive to the fluid pressure in the feed line 19s when it is greater than a given value and supplies a portion of the working fluid to the discharge line to maintain the pressure in the feed line 19s below the given pressure value.

On the other hand, a check valve 207 is arranged in the drain line 19d. The check valve 207 is also connected to the feed line 19s upstream of the one-way check valve 204 via a pilot line 208 and receives the pressure in the feed line therefrom as a pilot pressure. The check valve 207 is designed to be kept in an open position as long as the pilot pressure to be introduced through the pilot line 208 is maintained at a pressure level that is greater than or equal to a given pressure level. In the open position, the check valve maintains fluid communication between its inlet and outlet sides so that the working fluid in the drain line 19d can flow through it to the reservoir tank 21. On the other hand, the check valve 207 responds to the pilot pressure having fallen below the given pressure level and switches to a closed position. In the closed position, the shut-off valve blocks the fluid connection between the drain opening 18c and the reservoir container 21.

A pressure relief valve 209 is provided in parallel with the shut-off valve. The pressure relief valve 209 is arranged in a bypass line 210 connecting the upstream and downstream sides of the shut-off valve 207. The pressure relief valve 209 is normally held in a closed position to block the fluid communication therethrough. On the other hand, the pressure relief valve 209 is responsive to a fluid pressure in the drain line 19d on its upstream side which is higher than a set pressure, e.g., 30 kgf/cm, to establish fluid communication between the upstream and downstream sides of the shut-off valve so that the excessive pressure on the upstream side of the drain line 19d can be released therethrough. Therefore, the pressure relief valve 209 limits the maximum pressure at the set pressure. The set pressure of the pressure relief valve 209 corresponds to a specified offset pressure.

An oil cooler 211 is arranged in the drain line 19d for cooling the working fluid returning to the reservoir tank 21.

The following describes the supply of pressurized fluid using the pressure source unit described above.

While the automotive engine 200 is running, the fluid pump as the pressure unit 20 is driven. Therefore, the working fluid in the reservoir tank 21 is sucked through the suction line 201 and pressurized by the pressure unit 20. The pressurized working fluid is discharged through the outlet of the pressure unit 20 and supplied to the pressure control valve 18 through the feed line 19s including the pressure regulating orifice 202 and the one-way check valve 204. When the pressure control valve 18 is in a position of Fig. 3(A), the pressurized working fluid passes through the pressure control valve and enters the upper fluid chamber 15d of the hydraulic cylinder 15. On the other hand, when the pressure control valve 18 is in the position of Fig. 3(B) to block the communication between the feed line 19s and the upper fluid chamber 15d, the line pressure in the feed line increases. When the line pressure in the feed line 19s becomes greater than a set pressure of the pressure relief valve 206 in the relief line 205, the excess pressure higher than the set pressure is supplied to the drain line 19d via the pressure relief valve 206 and thereby returned to the reservoir tank 21.

The fluid pressure in the feed line 19s is also supplied to the shut-off valve 207 via the pilot line 208. As described above, the shut-off valve 207 is in the open position as long as the pilot pressure introduced through the pilot line 208 is maintained equal to or greater than the set pressure of the shut-off valve. Therefore, the fluid communication between the pressure relief valve 18 and the reservoir tank 21 is maintained. Therefore, when the pressure relief valve 18 is in the position of Fig. 3(B), the working fluid is returned to the reservoir tank 21 through the drain line 19d via the shut-off valve 207 and the oil cooler 211.

Since the shut-off valve 207 itself serves as a resistance to the fluid flow in the open position, the fluid pressure in the drain line 19d upstream of the controllable check valve 207 becomes excessively high, ie higher than the offset pressure P₀. Then, the pressure relief valve 209 becomes active and opens so that the excessive pressure of the working fluid can flow through the bypass line 210.

When the engine 200 stops, the pressure unit 20 stops operating. By stopping the pressure unit 20, the working fluid pressure in the feed line 19s drops. According to the pressure drop in the feed line 19s, the pilot pressure to be supplied to the check valve 207 via the pilot line 208 drops. When the pilot line 208 drops below or to the set pressure, the check valve 207 is switched to the closed position to block the fluid communication therethrough. As a result, the fluid pressure in the drain line 19d upstream of the check valve 207 becomes equal to the pressure in the upper fluid chamber 15d. Therefore, even if the working fluid leaks through a gap between the spool 19 and the inner peripheral surface of the valve bore 18a, it will not affect the fluid pressure in the upper fluid chamber 15d.

This is advantageous for maintaining the suspension characteristics of the suspension system regardless of the machine driving condition.

The microprocessor-based control unit 100 is provided for performing a control for adjusting the suspension characteristic depending on the vehicle running condition. The suspension control system including the control unit 100 is shown in Fig. 6. The control unit 100 shown in Fig. 6 includes a microprocessor 101 which comprises an arithmetic circuit 102, a memory 104 and an input/output unit 106. A vehicle speed sensor 108 is provided for monitoring the vehicle speed and generating a signal DV representative of the vehicle speed. The vehicle speed sensor 108 used in the shown embodiment of the suspension control system is designed to generate a pulse train having a frequency proportional to the vehicle speed as the signal DV representative of the vehicle speed. A lateral acceleration sensor 110 is also provided for monitoring lateral accelerations exerted on the vehicle and for generating a signal Gy representative of the lateral acceleration. The control system also includes a Longitudinal acceleration sensor 112 for monitoring longitudinal accelerations applied to the vehicle and generating a signal Gx representative of the longitudinal acceleration. In addition, vertical acceleration sensors 114FL, 114FR, 114RL and 114RR are provided for monitoring vertical accelerations on each of the front left, front right, rear left and rear right suspension systems 11FL, 11FR, 11RL and 11RR and generating signals representative of the vertical acceleration. Hereinafter, the acceleration sensor 114FL which monitors the vertical acceleration on the front left suspension system 11FL will be referred to as "FL vertical acceleration sensor". Similarly, the acceleration sensor 114FR which monitors the vertical acceleration on the front right suspension system 11FR will be referred to as "FR vertical acceleration sensor"; the acceleration sensor 114RL that monitors the vertical acceleration on the rear left suspension system 11RL will be referred to as "RL vertical acceleration sensor"; and the acceleration sensor 114RR that monitors the vertical acceleration on the rear right suspension system 11RR will be referred to as "RR vertical acceleration sensor". The signals representing the vertical acceleration generated by the respective FL, FR, RL and RR vertical acceleration sensors 114FL, 114FR, 114RL and 114RR will hereinafter be referred to as "FL vertical acceleration signal GzFL", "FR vertical acceleration signal GzRL", "RL vertical acceleration signal GzRL" and "RR vertical acceleration signal GzRR", respectively.

The vertical acceleration sensors 114FR, 114FR, 114RL and 114RR may in practice comprise a strain gauge or piezoelectric sensor mounted on the upper end portion of the suspension strut on the front left, front right, rear left and rear right wheels, respectively. The practical construction of the vertical acceleration sensor is disclosed in copending U.S. Patent Application 120,964, filed on November 16, 1987. The contents of this U.S. Patent Application 120,964 are incorporated by reference to supplement the present description.

The lateral acceleration sensor 110 is connected to a gain-controlled amplifier 116. The gain-controlled amplifier 116 controlled amplifier 116 is also connected to the input/output unit 106 of the microprocessor 101 through a digital-to-analog (D/A) converter 118 and receives a gain control signal R therefrom. The longitudinal acceleration sensor 112 is connected to gain controlled amplifiers 120F and 120R.

The FL, FR, RL and RRi vertical acceleration sensors 114FL, 114FR, 114RL and 114RR are connected to integrators 122FL, 122FR, 122RL and 122RR, respectively. Each of the integrators 122FL, 122FR, 122RL and 122RR integrates the corresponding FL vertical acceleration signal GzFL, FR vertical acceleration signal GzFR, RL vertical acceleration signal GzFR and RR vertical acceleration signal GzRR, respectively. The integrators 122FL, 122FR, 122RL and 122RR are connected to fixed gain amplifiers 124FL, 124FR, 124RL and 124RR. The amplifiers 124FL, 124FR, 124RL and 124RR are connected to inverted input terminals of adders 126FL, 126FR, 126RL and 126RR. The adders 126FL and 126FR have another inverted input terminal connected to the amplifiers 120F and 120R. On the other hand, the adders 126RL and 126RR have non-inverted input terminals connected to the amplifier 120. The outputs of the adders 126FL, 126FR, 126RL and 126RR are connected to non-inverted input terminals of further adders 128FL, 128FR, 128RL and 128RR. The adders 128FL and 128RR have inverted input terminals connected to the gain-controlled amplifier 116. On the other hand, the adders 128FR and 128RL have non-inverted input terminals connected to the gain-controlled amplifier 116. The output terminals of the adders 128FL, 128FR, 128RL and 128RR are connected to the metering solenoids 22 in the pressure control valves 18. Therefore, the electromagnets 22 adjust the position of the pistons 15c, and thereby the pressure in the upper fluid chambers 15d of the respective hydraulic cylinders 15A in the respective suspension systems 11FL, 11FR, 11RL and 11RR according to the outputs of the adders 128FL, 128FR, 128RL and 128RR, which serve as the suspension control signals SFL, SFR, SRL and SRR. The control signals output from the adder 128FL are hereinafter referred to as "FL control signal SFL"; the control signals output from the adder 128FR are hereinafter referred to as "FL control signal SFL" referred to as "FR control signal SFR"; the control signals output from the adder 128RL are hereinafter referred to as "RL control signal SRL", and the control signals output from the adder 128RR are hereinafter referred to as "RR control signal SRR".

The vertical acceleration signals GzFL, GzFR, GzRL and GzRR are used for compression control in which a relative displacement of the vehicle body and the suspension members is monitored. The suspension characteristics on the front left, front right, rear left and rear right suspension systems are respectively adjusted to suppress dive. For this purpose, the integrators 122FL, 122FR, 122RL and 122RR integrate the values of the vertical acceleration signals of the corresponding vertical acceleration sensors 114FL, 114FR, 114RL and 114RR and output the values representative of the relative displacement of the vehicle body and the suspension members in the compression and rebound directions from a predetermined initial position.

The lateral acceleration sensor 110, the gain-controlled amplifier 116 and the adders 128FL, 128FR, 128RL and 128RR introduce the roll suppression factor into the suspension control signals

SFL, SFR, SRL and SRR. Namely, in general, the gain controlled amplifier 116, the front left and rear right suspension systems are controlled in opposition to the front right and rear left suspension systems for suppressing the vehicle roll.

The gain of the gain-controlled amplifier 116 is adjusted by the gain control signal R from the microprocessor 101. The gain control signal R may be a voltage signal having a voltage level representative of a desired gain in the gain-controlled amplifier 116. The gain-controlled amplifier 116 of the embodiment shown is provided with a gain change characteristic so that the gain Ky is adjusted according to of the change in the gain control signal value R is varied linearly as shown in Fig. 7. Namely, as is apparent from Fig. 7, the gain Ky is a function (a R: a is constant) of the gain control signal value R. On the other hand, the gain control signal value R is given by a table read from the vehicle speed V. The change characteristic of the gain control signal value R is determined experimentally and is designed to avoid an unstable region in which a larger gain tends to cause self-induced sideways vibrations as shown in Figs. 8 and 9. The saturation point V₀ beyond which the change characteristic of the gain control signal value R changes is determined experimentally. In practice, the saturation point V₀ is set at about 10 km/h.

On the other hand, the longitudinal acceleration sensor 112, the gain-controlled amplifiers 120F and 120R, and the adders 126FL, 126FR, 126RL, and 126RR are arranged so that a pitch suppression factor is introduced into the suspension control signals SFL, SFR, SRL, and SRR to be output to the actuators. The gains of the gain-controlled amplifiers 120F and 120R are adjusted by gain control signals Ef and Er generated by the microprocessor 101 and supplied to the corresponding amplifier via D/A converters 130F and 130R, respectively. Since the pitch suppressing suspension control requires opposite control of the suspension characteristics on the front and rear suspension systems, the polarities of the inputs from the amplifiers 120F and 120R to the adders 126FL and 126FR are set different from those of the adders 126RL and 126RR. Therefore, when pitch suppressing suspension control takes place to stiffen or harden the front suspension systems, the rear suspension systems are controlled to make their suspension characteristics softer, and when the front suspension systems are made softer in the pitch suppressing control, the rear suspension systems are made harder.

The gains Kf and Kr of the gain-controlled amplifiers 120F and 120R are variable according to the linear characteristic shown in Fig. 10 depending on the gain control signals Ef and Er input from the microprocessor 101 via the D/A converters 130F and 130R. In the embodiment shown, the gain control signals Ef and Er are variable in value depending on the longitudinal acceleration Gx measured by the longitudinal acceleration sensor 112. Namely, when the longitudinal acceleration Gx is increased above zero, the gain control signal values Ef and Er increase with a given magnitude as shown in Fig. 11(a). In accordance with the change in the gain control signal values Ef and Er, the gains Kf and Kr of the gain controlled amplifiers 120F and 120R change, respectively, as shown in Fig. 11(b).

By controlling the gains Kf and Kr as described above and shown in Fig. 11(b), during a bow diving motion in which the longitudinal acceleration is positive, the amount of the gain Kf that determines the pitch suppressing suspension characteristic in the front suspension systems becomes larger than the amount of the gain Kr that determines the pitch suppressing suspension characteristic in the rear suspension systems. On the other hand, during the countermovement after bow diving, the amount of the gain Kr becomes larger than the amount of the gain Kf. This balances the amount of the pitching motion between the front and rear suspension systems. Therefore, suppression of bow diving can be successfully achieved.

On the other hand, the change characteristic of the gains Kf and Kr may be changed as shown in Fig. 12. As shown in Fig. 12, the gain change characteristic is reversed compared to that of Fig. 11(b). Such a change characteristic can be adapted to a vehicle having a change characteristic that causes a considerable acceleration surge. The change characteristics of the gains of the gain-controlled amplifiers 120F and 120R are not limited to the characteristics shown and can be set depending on the vehicle suspension characteristic or attitude change tendency.

In the following, the basic idea of the present invention of a roll suppressing suspension control with avoidance of the self-induced lateral vibrations is discussed based on the model shown in Fig. 12. In the model shown, a vehicle body 80 is suspended as a sprung mass on wishbones 82 which constitute an unsprung mass together with the vehicle wheel and the suspension unit. The lateral acceleration sensor 110 is mounted at an appropriate location on the vehicle body. The vertical acceleration sensors 114L and 114R are mounted on the top of the corresponding suspension unit. A lateral component of the roll energy causing a roll motion of the vehicle is therefore monitored by the lateral acceleration sensor 110. On the other hand, a vertical component of the roll energy is monitored by the vertical acceleration sensors 114L and 114R. The lateral acceleration signal of the lateral acceleration sensor 110 representative of the lateral component of the roll energy is amplified with a roll control gain Ky. On the other hand, the vertical acceleration values of the vertical acceleration sensors 114L and 114R representing the vertical component of the roll energy are amplified by a compression control gain K3. The amplified lateral acceleration and vertical acceleration signals are added to obtain roll suppressing suspension control signals SL and SR. The values of the roll suppressing control signals SL and SR are supplied to a hydraulic system H(s), ie, the actuator 22.

It is assumed that the conditions of the model of Fig. 13 are as follows:

[KEY DATA]

M is a weight of the sprung mass;

J is a roll moment of inertia;

K is a spring constant of the suspension;

L is a track width;

Hr is a roll center height;

Hg is a center of gravity height;

KL is a lateral strength of a tire;

KV is a vertical strength of the tire; C is an equivalent absorption of the cornering force of the wheel, which equivalent absorption represents the cornering force of the wheel and therefore is variable depending on the vehicle speed;

[CONSTANT]

H�1 is a height of the lateral acceleration sensor.

[VARIABLES]

Rys is a roll angle of the sprung mass;

Ryu is a roll angle of the unsprung mass;

x is a vertical displacement of the sprung mass;

y is a lateral displacement of the sprung mass;

z is a vertical displacement of the unsprung mass;

w is a displacement of an attack point.

The cornering force C of the wheel can be represented by:

C= RSL (V₁/V₂)

wherein

RSL is a glide angle;

V₁ is a lateral movement velocity;

V₂ is a longitudinal rotation velocity.

The value C becomes infinite (RSL·0/0) while the vehicle is stationary. The cornering force decreases as the vehicle speed increases from infinity. Therefore, in a fairly low vehicle speed region, the cornering force of the tire is relatively large. Therefore, in the fairly low vehicle speed region, the vehicle is laterally supported only by the cornering force C of the tire. In this vehicle speed region, if the roll suppressing gain K1 is relatively large, it tends to destroy the left-right balance of forces so that self-induced sideways vibrations are generated in response to relatively weak sideways accelerations. This means that in the suspension state of the vehicle with respect to roll stability, the stability factor becomes smaller as the roll suppressing gain K1 increases. This unstable region is shown in Fig. 14.

On the other hand, since the lateral guiding force C decreases with increasing vehicle speed, the vehicle is increasingly supported in the lateral direction by a series connection of the reduced lateral guiding force C and the lateral strength KL of the tire. Therefore, the stability factor increases as shown in Fig. 14.

In a state where self-induced lateral vibration occurs, a vibration frequency transmission coefficient for one vibration period at a vehicle speed of 5 km/h and 10 km/h is recorded. The measurement result is shown in Fig. 15. The lateral suspension system is unstable at the vehicle speed of 5 km/h and becomes stable at the vehicle speed of 10 km/h.

In the following, the practical suspension control operation for performing roll suppressing or anti-roll suspension control is discussed with reference to Fig. 16.

The program shown is executed at predetermined regular time intervals, e.g. every 20 ms. Immediately after the execution of the program begins, the signal DV representative of the vehicle speed is read and counted for a given period of time in a step 1002. The arithmetic circuit 102 then calculates the vehicle speed value V on the basis of the counted number of pulses in the given period of time in a step 1004.

Alternatively, the vehicle speed can also be derived by measuring the distance between the pulses, which is inversely proportional to the vehicle speed.

In a step 1006, the vehicle speed value V determined in the step 1004 is compared with a comparison value corresponding to the predetermined vehicle speed V₀ beyond which the change characteristic of the gain of the gain-controlled amplifier 116 changes. If the vehicle speed value V is greater than or equal to the comparison value representing V₀, in a step 1008 the gain control signal value R is set to the value R₀, which is the constant value used as long as the Vehicle speed is greater than the predetermined speed V�0;.

On the other hand, if the vehicle speed value V is smaller than the comparison value representing V₀ as checked in step 1004, the gain control signal value R corresponding to the vehicle speed is determined in step 1010. The control signal value R determined in step 1010 is smaller than the gain control signal value R₀ determined in step 1008. Namely, as can be seen from Fig. 8, the gain control value R determined in step 1010 changes linearly within a range between zero and R₀ in accordance with the vehicle speed in the range from zero to V₀.

After one of steps 1008 and 1010, in a step 1012 a gain control signal R is output to the gain-controlled amplifier 116.

Therefore, the response characteristic of the anti-roll suspension control in a region of fairly low vehicle speeds is considerably lower than that in a region of relatively high vehicle speeds. This successfully suppresses self-induced lateral vibrations of the vehicle and ensures ride comfort in a region of low vehicle speeds. On the other hand, since the self-induced lateral vibrations can thus be successfully avoided, a sufficiently high response characteristic of the anti-roll suspension control can be provided in a region of relatively high vehicle speeds. Therefore, the embodiment of the suspension control system shown can provide fairly high roll stability without deteriorating ride comfort in a region of fairly low vehicle speeds.

On the other hand, according to the program of Fig. 17, adjustment of the pitch suppression suspension control is carried out. The program of Fig. 17 is also triggered at given constant intervals, e.g., 20 ms.

Immediately after starting the execution, the signal value Gx representative of the longitudinal acceleration is read in a step 1102. In a step 1104, a longitudinal acceleration value GLOG is determined from the read signal value Gx representative of the longitudinal acceleration. In a step 1106, the gain control signal values Ef and Er are determined from the longitudinal acceleration value GLOG determined in the step 1104 in accordance with the characteristics shown in Fig. 11(a) and 11(b). In a step 1108, the gain control signals Ef and Er determined in the step 1106 are output to the gain-controlled amplifiers 120F and 120R.

Since the characteristics of the pitch suppressing suspension control can be adjusted to suit the suspension characteristics of the vehicle by setting the change characteristics of the gains of the gain controlled amplifiers 120F and 120R, an appropriate attitude control for the vehicle suspension system can be provided.

Therefore, the present invention achieves all of the objects and advantages sought.

Although the present invention has been disclosed in terms of the preferred embodiment to facilitate a better understanding of the invention, the invention can be embodied in various ways without departing from the principle of the invention. Therefore, the invention should be understood to include all possible embodiments and modifications of the embodiments shown which can be embodied without departing from the principle of the invention as set out in the appended claims. Namely, although the above discussion is directed to a hydraulic cylinder and a hydraulic circuit for generating a damping force for controlling the suspension characteristic in the actively controlled suspension system according to the invention, it should be possible to apply the same logic of roll suppressing and/or pitch suppressing suspension control to hydropneumatic suspensions, pneumatic suspensions, etc. Moreover, although the shown embodiment discloses the suspension control system that performs both roll suppressing and pitch suppressing control, the suspension system may be designed to perform only roll suppressing control.

Claims (9)

1. Anti-roll suspension control system for motor vehicles, with - a suspension system (13, 15) between a vehicle body (12) and a suspension member (13) rotatably supporting a vehicle wheel (14), which suspension system comprises a variable pressure pressure chamber (15d) filled with a working fluid of controlled pressure; - a pressure adjustment device (18, 20) in connection with the pressure chamber with variable pressure for adjusting the pressure of the working fluid in the pressure chamber; - a sensor (112) for monitoring the roll conditions of the vehicle and for generating a signal representative of the roll intensity; - a control unit (100) which receives the signal representative of the roll intensity, generates an anti-roll suspension control signal on the basis of the signal representative of the roll intensity, and controls the pressure adjusting device (18) for adjusting the suspension characteristics for suppressing roll of the vehicle; and - a vehicle speed sensor (108) which generates a signal representative of the vehicle speed, characterized in that - the control unit (100) receives the signal representative of the vehicle speed and adjusts the response characteristics of the anti-roll suspension control in accordance with the signal representative of the roll intensity in dependence on the signal value representative of the vehicle speed, so that the response characteristic is adjusted to be lower than a predetermined first value according to a predetermined response characteristic curve which is substantially proportional to the signal representative of the vehicle speed, so that self-induced vibrations are suppressed which occur due to the anti-roll suspension control in a roll direction while the signal value representative of the vehicle speed is lower than a predetermined vehicle speed and the response characteristic is maintained at or near the first value so that the anti-roll effect is enhanced while the signal value representative of the vehicle speed is greater than or equal to the predetermined vehicle speed. 2. Anti-roll suspension control system according to claim 1, characterized in that the control unit (100) changes the response characteristic between the first value and the value zero according to the signal value representative of the vehicle speed while the signal value representative of the vehicle speed is lower than the given value. 3. Anti-roll suspension control system according to claim 1, characterized in that the control unit (100) is set such that the given value is a value representative of a substantially low speed. 4. Anti-roll suspension control system according to claim 1, characterized in that the pressure adjusting means comprises a pressure fluid source (20) connected to the variable pressure pressure chamber (15d) and a pressure control valve (18) including an electrically operable actuator (22b), which actuator is connected to the control unit and receives the anti-roll suspension control signal and operates the pressure control valve so as to adjust the amount of working fluid introduced and discharged with respect to the variable pressure pressure chamber, and thus adjust the damping characteristic of the suspension system. 5. Anti-roll suspension control system according to claim 4, characterized in that the control unit (100) comprises a gain-controlled amplifier (116) which determines the response characteristic, which control unit adjusts the gain of the amplifier in accordance with the signal representative of the vehicle speed. 6. Anti-roll suspension control system according to claim 5, characterized in that the control unit (100) compares the signal representative of the vehicle speed with a predetermined low speed criterion and sets the gain of the gain-controlled amplifier (116) to a constant first value if the signal value representative of the vehicle speed is greater than or equal to the low vehicle speed criterion, and lowers the gain in accordance with the signal value representative of the vehicle speed if it is less than the low vehicle speed criterion. 7. Anti-roll suspension control system according to claim 6, characterized in that the control unit (100) changes the gain of the gain-controlled amplifier (116) in a range between the first value and the value zero depending on the signal value representative of the vehicle speed. 8. Anti-roll suspension control system according to claim 7, characterized in that the control unit (100) changes the gain of the gain-controlled amplifier (116) linearly according to the signal value representative of the vehicle speed. 9. Anti-roll suspension control system according to one of claims 1 to 8, characterized by a third sensor in the vicinity of the suspension system for monitoring the vertical movement of the vehicle and generating a signal representative of the vertical movement of the body of the vehicle, which control unit (100) receives the signal representative of the magnitude of the vertical movement and generates an anti-roll suspension control signal on the basis of the signal representative of the vertical movement magnitude and the signal representative of the roll magnitude.